Axial-and radial-flow, multistage centrifugal compressor



Feb. 14, 1967 J. RURIK ETAL 3,303,989

AXIAL-AND RADIAIrFLOW, MULTIS'IAGE CENTRIFUGAL COMPRESSOR Iiled March1964 4' Sheets-Sheet 1 I v'l'i l JOSEF RUR/K HELLMUT WE/A/R/CH Feb. 14,197 J. RURIK ETAL. 3,303,989

AXIAL-ANDRADIAb-FLOW, MULTISTAGE CENTRIFUGAL COMPRESSOR Filed March 2,.1964 4 Sheets-Sheet 2 IHHHI A .T TORNEYS Feb. 14, 1967 J. RURIK ETALAXIAUAND RADIAL-FLOW, MULTISTAGE CBNTRIFUGAL COMPRESSOR Filed March 2,1964 4 Sheets-Sheet 5 FlG.3

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Feb. 14, 1967 Filed March 2, 1964 P9553 ups k/ss J. RURIK ETAL 3,303,989AXIAL-AND RADIAL-FLOW, MULTISTAGE CENTRIFUGAL COMPRESSOR 4 Sheets-Sheet4 w l l VOLUMETFP/C INTAKE F l G. 5

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United States Patent 3,303,989 AXIAL- AND RADIAL-FLOW, MULTISTAGECENTRIFUGAL COMPRESSOR Josef Rurik, Oberhausen-Sterkrade, and HellmutWeinrich, Heidenheim (Brenz), Germany, assignors to GutehofinungsllutteSterkrade Aktiengesellschaft, Oberhausen-Sterkrade, Germany Filed Mar.2, 1964, Ser. No. 348,729 Claims priority, application Germany, Mar. 2,1963, G 37,187 4 Claims. (Cl. 23045) Purely axial-flow and purelyradial-flow compressors are known. The axial-flow compressor hasgenerally a relatively high efliciency, particularly at relatively highpowers. This higher efliciency is particularly obtained in thepartial-load region by a suitable adjustment of the stationary blades.On the other hand, axial-flow compressors are, owing to design reasons,less suitable for interstage cooling so that higher pressure ratiosrequire a correspondingly large number of stages. Inverse conditions areobtained with radial-flow compressors. They have generally a relativelylower efficiency, particularly in the partial-load region, because theswirl control .required for this purpose requires the incorporation ofspecial stationary blades before the moving blades. On the other hand,interstage cooling can be provided for without special difiiculties inradial-flow compressors.

Apart from the use of blades having an axial-flow inlet portion and aradial-flow outlet portion, it has already been proposed to arrangeaxial-flow blades and radialflow blades in different forms in amultistage compressor. It has not yet been possible, however, to obtainan optimum utilization of the properties of both types of blades.

For the practical success of such a combined axialand radial-flowcompressor, its performance under different operating conditions is ofgreat importance. Hence, it is an object of the invention to provide asolution in this respect mainly for compressors which in spite of aconstant operating speed imparted to the compressor by an electricmot-or drive are required to have not only a high overall efficiency butalso an optimum power consumption at partial loads, at a capacity of,e.g., 70% of the normal full-load capacity.

Hence, the invention utilizes adjustable stationary blades, such as areknown per se in axial-flow stages. In known combined axialandradial-flow compressors, such adjustable stationary blades have beenused in all axialfiow stages, and the resulting greater length of thecompressor has partly been compensated by the use of at least oneradial-flow impeller as the last stage. The invention, however, is basedon a recognition which is of special significance for relatively highcompression ratios, i.e., for compressors having a large number ofstages. This recognition resides in that the desired object can be quiteadequately accomplished even when adjustable stationary blades areprovided only to a limited extent. Hence, the invention resides in thatonly a part of the axial-flow stages is provided with adjustablestationary blades and at least one axial-flow stage havingnon-adjustable stationary blades is disposed between these adjustableaxialfiow stages and the succeeding radial-flow stage or stages.

If a plurality of successive axial-flow stages .are to be provided withadjustable stationary blades, it is advantageous to use for this purposea common adjusting linkage. Intermediate levers of different length maybe provided "ice in known manner for an adjustment through differentangles. In this connection, a further feature of the invention residesin that the length of the intermediate levers is such that the largestangle of adjustment is obtained in the first stage and the angle ofadjustment obtained in each succeeding adjustable stage is somewhatsmaller than in the preceding stage in such a manner that a zeropressure rise is obtained in each of these stages when they are allcompletely shut off.

The above-mentioned features of the invention may be used with differentarrangements of the axial-flow and radial-flow stages. For instance, aplurality of axial-flow stages serving as a low-pressure section and aplurality of radial-flow stages serving as a succeeding high-pressuresection may be arranged on a common shaft, and a cooler may beinterposed in known manner between any two adjacent radial-flow stages.Alternatively, two stage groups, each of which comprises preferably aplurality of axial-flow stages and a succeeding radial-flow stage, maybe arranged for mutually opposed directions of axial flow in such amanner that the inlet axial-flow stages of the two groups lie on theoutside beside the shaft bearings and their outlet radial-flow stageslie on the inside one beside the other. A cooler is suitably interposedbetween the two combined axialand radial-flow stage groups.

In conjunction with the last-mentioned type, particularly highcompression ratios (up to 10 or higher) may be achieved if the secondcombined aXialand radial-flow stage group is succeeded by a secondcooler followed by a single radial-flow stage, which is preferablyfloatingly mounted on the shaft end opposite to the drive motor.

Several illustrative embodiments of the invention are shown partlydiagrammatically on the drawings.

FIG. 1 shows a compressor which comprises four lowpressure, axial-flowstages succeeded by three high-pressure, radial-flow stages.

FIG. 2 shows a compressor which comprises adjustable axial-flow stages811 and a non-adjustable axial-flow stage 16.

The compressor shown in FIG. 3 has two stage groups, the first of whichconsists of five series-connected axialflow stages and a succeedingradial-flow stage, whereas the second stage group has four axial-flowstages and a succeeding radial-flow stage.

FIG. 4 shows .a modification of FIG. 3 and includes an additionalradial-flow stage, which is floatingly mounted on one end of the shaftoutside the shaft bearing and succeeds the second stage group.

FIG. 5 is a graph showing an operating characteristic curve a for thepressure rise in percent of a centrifugal compressor plotted against thevolumetric intake rate.

FIG. 6 shows a comparison between the power consumption in percent of acooled radial-flow compressor having swirl-controlled first stages andof an axialand radial-flow compressor having adjustable stationaryblades in the axial-flow section, in the same range of about -110% ofthe normal volumetric intake rate at full load The compressor shown inFIG. 1 comprises a low-pressure section '1 comprising four axial-flowstages, and a succeeding high-pressure part '2 consisting of threeradialflow stages. All axial-flow stages'and radial-flow stages aremounted in a common housing 3 and are carried by a common shaft 4between two supporting bearings 5, 6 of this shaft. The last radial-flowstage has an inlet in a direction which is opposite to the two otherradialfiow stages and the axial-flow stages so that the axial thrust isat least partly compensated. More radial-flow stages may be reversed inthis way, except for the first one, the inlet of which is suitablydirectly opposite to the outlet of the last axial-flow stage, forflow-dynamical reasons.

An interstage cooler is arranged between every two adjacent radial-flowstages. In the present case, two interstage coolers are provided, whichare suitably symmetric-ally arranged on both sides of the compressor andfor this reason are not apparent from the longitudinal sectional viewwhich is shown.

The stationary blades of the axial stages are adjustable according tothe invention with the aid of a common adjusting linkage 7. Details ofsuch an adjusting device, which is known per so, are apparent on alarger scale from FIG. 2. A special feature resides in that theintermediate levers 12-15 leading to the stationary blades of thevarious adjustable axial-flow stages 8-11 have different lengths. Theintermediate lever 12 associated with the first axial-flow stage 8 isthe shortest. The intermediate lever of each succeeding, adjustableaxial-flow stage is somewhat longer than that of the preceding stage. Inspite of a joint operation of all adjustable stationary blades, theblades of succeeding stages are adjusted through different angle-s. Therelatively largest angle of adjustment is obtained in the first stage 8and the relatively smallest angle in the last stage 11. Tests have shownthat the different angles of adjustment are suitably selected so that areduced total pressure rise of all adjustable axial-flow stages isdivided so as to obtain progressively increasing stage pressure rises,and that for a zero overall pressure rise the pressure rise of eachstage is also zero.

Such an adjusting device enables the respective axialflow stages to beentirely or largely shut olf in consideration of. their different bladelengths so that during partial-load operation the fluid to be compressedcan flow substantially without obstruction through these stages with apower consumption which is very small in comparison with that of theswirl control employed with radialfiow stages.

In the example shown in FIG. 2, the adjustable axialflow stages 811 aresucceeded by-a non-adjustable axialflow stage 16.

According to FIG. 3, a shaft 19, which is driven by a clutch 18 from amotor 17, carries two stage groups 20, 21 between two shaft bearings 22,23. One stage group 20 comprises five series-connected axial-flow stagesand a succeeding radial-flow stage. The other stage group 21 comprisesfour axial-flow stages and a succeeding radial-flow stage. Both stagegroups are disposed in such a relation to each other that the inletaxial-flow stages are disposed on the outside beside the shaft bearingsand the radial-flow stages are arranged on the inside one beside theother. This results in a considerable compensation of the axial thrustas well as in a relatively low pressure differential at the two outerstulfing-boxes beside the shaft bearings and on the inner stuffing-boxbetween the two radial stages so that these stuffing boxes can easily besealed.

A cooler 24 is arranged between the two stage groups 20, 21. This coolermay be attached to the compressor housing, not shown, or may be separatefrom this hous ing. As a result of the effective interstage cooling, thevolume of the pro-compressed fluid is considerably reduced from theoutlet of the radial stage of the first stage group 20 to the inlet of.the first axial-flow stage of the second stage group 21.

- In FIG. 4, like reference characters designate parts which are similarto those used in FIG. 3. FIG. 4 shows an additional radial-flow stage25, which is floatingly mounted on one end portion of the shaft 19outside the shaft bearing 23 and succeeds the second stage group 21. Asecond cooler 26 is arranged between the outlet radial-flow stage of thesecond stage group 21 and the additional radial-flow stage 25. Whileretaining all other advantages of a compressor according to theinvention, this embodiment enables an even higher overall compressionratio.

In the embodiments shown in FIGS. 3 and 4, the axialflow stages areprovided at least in part with adjustable stationary blades for a goodadaptation to changing operating conditions.

In FIG. 5, curve a represents the operating characteristic curve of acentrifugal compressor. The pressure rise in percent is plotted againstthe volumetric intake rate.

FIG. 6 represents the power consumption in percent with the range ofabout 70=l-10% of the normal volumetric intake rate at full load for acooled radial-flow compressor which has swirl-controlled first stagesand an axialand radial-flow compressor having adjustable stationaryblades in the axial-flow section. These curves b (for the pureaxial-flow compressor) and c (for the combined axialand radial-flowcompressor) have been measured and computed for two compressors whichhave the same final pressure and indicate that the compressor accordingto the invention has a lower power consumption throughout the rangewhich is represented. This difference is particularly high in thepartial-load region.

What we claim is:

1. A multi-sta-ge centrifugal compressor comprising, in combination, acommon drive shaft extending between a pair of axially spaced bearings;two serially connected multi-stage sections supported on said commondrive shaft between said bearings; a common housing enclosing bothmulti-stage sections between said bearings; each multistage sectioncomprising plural axial flow fluid inlet stages and at least one radialflow fluid outlet stage, the axial flow stages being at the axiallyouter end of each section and the radial flow stages being disposedaxially between the two sets of axial flow stages; the fluid flowingserially through the two sect-ions in respective opposed axialdirections; at least one axial stage of each section having adjustableguide vanes; adjusting means operable to adjust said guide vanes; and acooler connected to the outlet of at least one radial stage.

2. A multi-stage centrifugal compressor, as claimed in claim 1, in whichplural axial stages of each section have said adjustable guide vanes;said adjusting means including respective crank arms each connected tothe guide vanes of a respective axial stage; a common operator connectedto the free ends of said crank arms to simultaneously pivot all of saidcrank arms; said crank arms having different lengths with the lengthsof. the crank arms varying in a manner such that the adjustment anglefor the first axial stage has the largest value, with the adjustmentangle for each succeeding axial stage being progressively smaller. I

3. A rn-ulti-stage centrifugal compressor, as claimed in claim 2, inwhich the unequal adjustment angles for the guide vane-s of respectiveaxial stages have values such that, with zero pressure increase for allof the axial stages of the respective section, the pressure increase foreach individual axial stage is equal to zero.

4. A m-ulti-sta-ge centrifugal compressor, as claimed in claim 1,including an individual radial flow stage supported on said common driveshaft and arranged, in the direction of fluid flow through saidcompressor, downstream of said two serial connected multi-stagesections; and a cooler connected between the outlet of the downstreamone of the two serially connected multi-stage sections and the inlet ofsaid individual radial stage.

(References on following page) References Cited by the Examiner UNITEDSTATES PATENTS Banner 230-130 Keller 230-119 Planiol 230--119 Hagen230114 Wild-e 230-114 6 FOREIGN PATENTS 3/ 1963 Great Britain. 12/ 1915Switzerland. r 12/ 1915 Switzerland. 71,244 12/ 1915 Switzerland.102,821 1/ 1924 Switzerland.

LAURENCE V. EFNER, Primary Examiner.

1. A MULTI-STAGE CENTRIFUGAL COMPRESSOR COMPRISING, IN COMBINATION, ACOMMON DRIVE SHAFT EXTENDING BETWEEN A PAIR OF AXIALLY SPACED BEARINGS;TWO SERIALLY CONNECTED MULTI-STAGE SECTIONS SUPPORTED ON SAID COMMONDRIVE SHAFT BETWEEN SAID BEARINGS; A COMMON HOUSING ENCLOSING BOTHMULTI-STAGE SECTIONS BETWEEN SAID BEARINGS; EACH MULTISTAGE SECTIONCOMPRISING PLURAL AXIAL FLOW FLUID INLET STAGES AND AT LEAST ONE RADIALFLOW FLUID OUTLET STAGE, THE AXIAL FLOW STAGES BEING AT THE AXIALLYOUTER END OF EACH SECTION AND THE RADIAL FLOW STAGES BEING DISPOSEDAXIALLY BETWEEN THE TWO SETS OF AXIAL FLOW STAGES; THE FLUID FLOWINGSERIALLY THROUGH THE TWO SECTIONS IN RESPECTIVE OPPOSED AXIALDIRECTIONS; AT LEAST ONE AXIAL STAGE OF EACH SECTION HAVING ADJUSTABLEGUIDE VANES; ADJUSTING MEANS OPERABLE TO ADJUST SAID GUIDE VANES; AND ACOOLER CONNECTED TO THE OUTLET OF AT LEAST ONE RADIAL STAGE.